Hybrid vehicle powertrain with a multiple-ratio power transmission mechanism

ABSTRACT

A control method and system for a hybrid vehicle powertrain with an electric motor in a power flow path between an internal combustion engine and a multiple-ratio geared transmission. Motor torque, which is added to engine torque to obtain an effective transmission input torque, is modulated to achieve smooth power-on upshifts and coasting downshifts.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims the benefit of U.S. provisional application Ser.No. 60/501,706, filed Sep. 10, 2003.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The invention relates to hybrid vehicle powertrains with an internalcombustion engine, a multiple-ratio transmission and an electricstarter/alternator motor between the engine and the transmission.

2. Background Art

A contemporary automotive vehicle powertrain typically includes ahydrokinetic torque converter disposed between a transmission withmultiple-ratio gearing and an internal combustion engine. The turbine ofthe torque converter transfers power to the power input element of themultiple-ratio gearing.

The presence of the torque converter in a vehicle powertrain of thistype introduces hydrokinetic power losses, particularly during vehiclestart up and advanced throttle downshifts. The power losses aremanifested by thermal energy build up in the hydrokinetic torqueconverter fluid, which requires a heat exchanger to maintain anacceptable fluid temperature. Attempts have been made to reduce powerlosses normally associated with torque converter automatic transmissionby eliminating the torque converter and replacing it with an electricmotor (starter/alternator). A powertrain configuration of this typetakes advantage of the performance of an internal combustion engine withthe advantages of an electric motor that complements the speed andtorque characteristics of the engine. It improves fuel economy of thepowertrain while reducing undesirable exhaust gas emissions. It alsopermits the engine to be deactivated when the vehicle is at rest. Themotor, which could be a high voltage induction motor, is available toprovide added performance. Further, the engine can be disconnected fromthe power flow path of the powertrain as the induction motor suppliesdriving torque. When the motor is not required for providing drivingtorque, it can function as an alternator.

Further, in a powertrain of this type, the kinetic energy stored in amoving vehicle can be collected by charging a high voltage batteryduring deceleration.

In a conventional powertrain with a hydrokinetic torque converter, shiftquality may be improved by controlling transmission input torque.Transmission input torque is reduced during a power on upshift byretarding the engine spark to reduce engine output torque. This improvesboth the durability of the oncoming friction element and the smoothnessof the upshift event. Torque modulation using spark retard will satisfythe timing and repeatability requirements to satisfy shift qualitytargets, but this wastes a small amount of energy during the shift thatheats the exhaust gases. Further, it can only reduce torque, notincrease it. Torque modulation also can be accomplished by using a fuelcut off to reduce engine torque, but restoring engine torque following ashift event often is not repeatable using fuel control.

In a conventional powertrain using a hydrokinetic torque converter, acoast mode occurs whenever the accelerator pedal is off, both with andwithout braking. As the vehicle slows, a coasting downshift must be doneto keep the engine speed within the desired range. In the case of acoasting downshift for a hybrid electric vehicle, the regenerativebraking function coincides with the coast mode. Since the motor islocated between the engine and the transmission, the coast downshift isdone with a significant level of negative torque at the input to thetransmission. This is an operating condition that differs from acondition found in a conventional powertrain, where coasting downshiftsare done with only a slight negative or positive torque at thetransmission input.

SUMMARY OF THE INVENTION

The invention includes a strategy for controlling power-on upshifts andcoasting, reduced-throttle downshifts in a multiple-ratio transmissionfor a converterless powertrain of the kind described in the precedingbackground discussion. The strategy uses a starter/alternator, sometimesreferred to herein as a motor, in conjunction with control oftransmission friction elements to provide smooth coasting downshiftsthat effect regenerative braking torque in the motor. The invention alsoincludes a strategy that uses starter/alternator torque for torquemodulation at the input of the transmission during an upshift.

The hybrid electric vehicle powertrain of the invention provides asignificant improvement in fuel economy without sacrificing convenienceand comfort associated with a conventional vehicle. The powertrainprovides full hybrid characteristics including improved engine stops andstarts, electric motor boost, regenerative braking and electric drive.The key components of the powertrain are packaged within a basetransmission assembly wherein the torque converter of the basetransmission assembly is removed and replaced with an electric highvoltage motor that serves as a starter, an alternator and a hybridtraction motor. A clutch may be added between the motor and the engineto allow full electric drive capability. Transmission fluid is used tocool the motor, and the added clutch may be controlled with atraditional transmission electro-hydraulic system.

The invention provides active control of the transmission frictionelements and coordinates that control with control of the motor. Thisresults in torque modulation of shift events by reducing transmissioninput torque during a shift event.

During a power on upshift, the transmission ratio is changed to a ratiowith lower torque multiplication. When the powertrain is producingpositive torque at the wheels, the engine operating speed is lowered.This upshift normally is commanded by the strategy. It is not a reactionto a driver demand. Thus, high shift quality has added importance.

To improve shift quality, the change in the magnitude of vehicleacceleration and the derivative of the vehicle acceleration experiencedduring the shift is reduced. To achieve this reduction in accelerationchange, the output torque should be as smooth as possible. During theshift event, the output torque is a function of several variables. Twovariables that are of importance are the input torque and the torquecapacity of the controlling element. The torque capacity of thecontrolling element, in turn, is a function of hydraulic pressure. Thisvariable will be described first.

The input torque for the hybrid electric vehicle powertrain is thealgebraic sum of the engine torque and the motor torque. A combinationof hydraulic pressure control and motor torque control is used to smoothand shape the output torque during power-on upshifts. The shaping of theoutput torque reduces occurrences and severity of torque fluctuations ortorque shuffle at the output shaft, which normally would be associatedwith elimination of the torque converter.

To schedule the pressure and torque computations and the outputcommands, a power-on upshift is divided into five modes. The first mode(Mode 0) is an initialization mode, which is called upon when thedesired gear is not the same as the current gear. It is used toinitialize variables for the start of the upshift. Mode 1 is thenentered during the next mid-ground loop execution of a transmissionmicroprocessor controller. This mode lasts for a predetermined fixedtime. It is used to prepare the friction elements for pressure control.Mode 2, which is the torque phase of the shift, is entered after a fixedtime for Mode 1 execution has expired. In Mode 2, the oncoming controlpressure is commanded to ramp from a pressure that is a function ofinput torque and speed to an initial shift pressure. This pressure rampis used to control the rate of pressure increase for the oncomingfriction element. Triggers are monitored to indicate when the oncomingelement has sufficient capacity to allow the release of the off-goingelement. At this point, the off-going friction element pressure iscommanded to a predetermined clutch stroke value. Mode 2 is completeafter a percent of ratio change completion has exceeded a predeterminedtrigger value or a set time has expired.

Mode 3, which is the inertia phase of the shift, involves a majorportion of the speed ratio change interval. In Mode 3, a PID controller,based on a comparison of the desired oncoming slip speed and the actualoncoming slip speed, generates an oncoming friction element pressurecommand. This closed loop pressure control is calibrated so that theshift rate is initially high. It then is decreased as the shift pressureis reduced during the inertia phase. The reduced torque capacity and theresulting reduced deceleration is used to reduce the severity of an endshift shock that can cause torque shuffle after the shift is completed.

Mode 4 is entered when the percent shift complete exceeds apredetermined value. In Mode 4, the pressure is increased to the valuerequired for non-shifting operation.

The other variable that is controlled during a power-on upshift, asmentioned above, is the transmission input torque. In a conventionalautomatic transmission, input torque is reduced during a power-onupshift, as previously mentioned, by retarding the engine spark toreduce engine output torque.

The motor in the powertrain of the invention has the ability to providesignificant positive torque and regenerative torque with a relativelyquick response time. Motor control strategy is implemented using the 5shift modes corresponding to the shift modes for the pressure control.In Mode 1, the maximum value of torque the motor will be commanded toabsorb is determined as a function of pedal command. In Mode 2, themotor is commanded to start battery charging after start of the speedchange is sensed. The input torque reduction is commanded at an optimumtime in the shift interval to avoid reducing the minimum torque duringMode 2, which would reduce shift quality.

A shaping function, stored in microprocessor memory, is used todetermine the actual level of the torque that the motor is commanded toabsorb. In Mode 3, as the shift progresses, the motor is commanded toabsorb less torque (i.e., to supply less negative torque) therebyrestoring the transmission input torque to its unmodulated value by theend of the shift. The increasing input torque reduces the shift rate atthe end of the shift, thereby further mitigating the torque shuffleproblem at the torque output shaft of the transmission.

In the case of a coasting downshift, the regenerative braking functioncoincides with the transmission coast mode. Because the motor is locatedin the powertrain between the engine and the transmission, the coastmode has a significant level of negative torque at the input for thetransmission. From a regenerative braking perspective, it is desirableto have the transmission solidly in gear so that the maximum amount ofenergy can be collected. This requires the transmission to complete theshift in a short period of time. Also, because of the regenerativebraking, the coasting downshift is done with a variable and significantnegative input torque. The requirements for short shift times and thehigh level of negative torque delivered to the transmission cause theshift to be more of a challenge than a coasting downshift in aconventional powertrain.

The smooth coasting downshifts maintain regenerative braking torque inthe motor in a manner that resembles a so-called “mirror image” of apower-on upshift. The 5 modes of a power-on upshift, previouslydescribed, are used also in a coasting downshift.

The strategy for a coasting downshift requires a desired shift time thatis dependent on the rate at which the vehicle is decelerating. Fastershifts are required during fast braking rates in order to avoid acondition in which one shift is stacked on top of another. Further, therequired change in engine speed depends on the vehicle decelerationrate. For a coasting downshift, the change in vehicle speed during theshift can be significant under high vehicle deceleration rates. Sincethe vehicle deceleration rates can be very different under differentoperating conditions, the strategy takes vehicle deceleration rate intoaccount.

The torque of the motor during a coasting downshift is increased beforethe start of the inertia phase to keep the transmission input fromdecelerating. It also is desirable to not completely bring the motortorque back to where it was at the beginning of the coasting downshiftas the vehicle system controller normally commands less regenerativetorque in lower gears (higher torque multiplication ratios). Less motortorque is desirable in lower gears since the transmission output torqueshould be consistent before and after the shift to maintain good brakingfeel.

As mentioned earlier, a coasting downshift event is characterized by 5different modes, as in the case of the power-on upshift. The first mode(Mode 0) is an initialization mode, which is called upon when thedesired gear is not the same as the current gear. It is used toinitialize variables for the start of the upshift. The coastingdownshift continues in Mode 1 with the oncoming friction element beingboosted to fill the friction element actuator quickly while theoff-going friction element pressure is set to a value just sufficient tohold input torque. In Mode 2, the oncoming element is set to its desiredvalue as it completes its stroke. The off-going element is ramped down.The off-going element holds the transmission in its current torque ratiowhile the oncoming element is stroking. Then the off-going element isreleased and the oncoming element can control the transmission inputspeed up to the new speed ratio.

The motor produces negative torque, but it is controlled to a lowerabsolute level during the shift in such a way that it acts in a fashionthat is the opposite of torque modulation during a power-on upshift. Thetorque is increased from its level before the shift to a predeterminedlevel during the shift through a ramp function. This ramp is started inMode 1 and continues until a desired torque value is reached at apredetermined value, or until the motor torque is requested to returnnear the completion of ratio change.

Once a predetermined percentage of a coasting downshift is completed,Mode 2 is complete. In Mode 3, which is the inertia phase, the oncomingpressure is commanded through a closed loop PID controller to follow aspeed profile to the downshifted gear. The output pressure command ofthe PID control is prevented from going below a minimum value (clip). Achange to an increased minimum clip on the pressure command makes theshift more aggressive under two conditions: First, for a manualdownshift it is desirable to have an aggressive feel so that the vehiclenoticeably slows in response to the driver's desire for hill braking;and secondly, under a fast vehicle deceleration rate, the shift iscompleted faster and a more aggressive shift rate is acceptable.

Motor torque is expected to be at the level desired to complete theratio change. As the shift is progressing toward completion, the motortorque is ramped back toward the original torque level that wascommanded at the start of the shift. The point at which the ramp isbegun is a function of percentage shift complete. To make the rampindependent of the vehicle deceleration rate, the ramp rate is also afunction of percentage shift complete. Mode 3 is exited when the shiftpercentage complete is near 100%.

In Mode 4, the oncoming pressure command is increased to a maximumcommand at the completion of the shift. Motor control is then returnedby the transmission controller to the vehicle system controller. Ifthere is a mismatch at that time between the motor torque command afterthe shift and the new desired regenerative braking torque command by thevehicle system controller, the command is filtered through a first orderfilter to the new level.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross-sectional view of a hybrid vehicle powertraincomprising an engine, a multiple-ratio transmission and an electricmotor, the motor being located between the transmission and the engine;

FIG. 1 a is a schematic diagram of the elements of the powertransmission mechanism of FIG. 1;

FIG. 1 b is a chart showing the clutch and brake engagement and releasepattern for each of four forward driving ratios and a reverse ratio forthe transmission of FIG. 1;

FIG. 2 is a partial cross-sectional view of a hybrid vehicle powertrainembodiment similar to the embodiment of FIG. 1, although in FIG. 2 aslipping wet clutch connection between the engine and the torque inputelement of the multiple-ratio transmission is lacking;

FIG. 3 is a schematic representation of an overall control system forcontrolling a hybrid powertrain of the type shown in FIGS. 1 and 2;

FIG. 4 a is a time plot for the speed ratio during a power-on upshiftevent;

FIG. 4 b is a time plot of the starter alternator torque during apower-on upshift event;

FIG. 4 c is a time plot of the torque at the powertrain output shaftduring a power-on upshift event;

FIG. 5 is a time plot of motor torque and the motor-speedcharacteristics for a high voltage induction motor of the kind adaptablefor use in the powertrain of the invention;

FIG. 6 is a time plot of simulated powertrain output torque, togetherwith a speed ratio plot during a power-on upshift, wherein torquefluctuations at the output shaft for the powertrain of the invention iscompared to a similar powertrain that does not include the controlstrategy of the invention;

FIG. 7 a is a time plot showing engine speed, output torque,transmission gear and shift mode during a power-on upshift event;

FIG. 7 b is a time plot for a power-on upshift, which shows variationsof engine torque, transmission input torque, transmission output torque,engine speed and motor torque for each of the shift modes;

FIG. 7 c is a time plot for the pressures on the friction elementsduring a power-on upshift event;

FIG. 8 a is a time plot of motor torque during a coasting downshift;

FIG. 8 b is a time plot of the pressure command on the off-going clutchand the oncoming clutch during a coasting downshift;

FIG. 8 c is a time plot of the input speed for the transmission during acoasting downshift;

FIG. 9 is a time plot of the output torque during a coasting downshift;

FIG. 10 is a time plot of the pressure on the oncoming clutch and theoff-going clutch during a coasting downshift; and

FIG. 11 is a time plot of transmission input speed during a coastingdownshift.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT(S)

In FIG. 1, reference numeral 10 designates schematically an internalcombustion engine for an automotive vehicle. Numeral 12 designatesgenerally a multiple-ratio automatic transmission.

The engine 10 includes a crankshaft 14 journalled at 16 in end wall 18of the engine housing. An intermediate housing 20 is located between theend wall 18 of the engine 10 and wall 22 for the transmission 12.

The intermediate housing encloses a stator 24 of an electric motor. Thestator and the stator windings are secured to an interior machinedsurface 26 of the housing 20. A rotor assembly 28 is situated within thestator and separated from the stator by an air gap designated by numeral30.

A wet clutch cylinder 32 is secured to the rotor assembly 28. A supportbearing shaft 34 rotatably supports the clutch cylinder 32 and issecured by bolts 36, or by other suitable fastening techniques, to thetransmission wall 22.

An annular piston 38 situated within the clutch cylinder 32 cooperateswith the cylinder to define a pressure chamber in communication withfluid pressure passage 40 in the support shaft 34. Passage 40communicates with passage structure extending to a control valve body,shown generally at 42, which is secured to the upper side oftransmission housing 44.

A slipping wet clutch disc assembly 46 has clutch plates secured tocylinder 32. Internally splined clutch discs are carried by clutchelement 48, which in turn is connected, preferably by drive splines, tothe hub 50 of a spring damper assembly 52. A damper drive plate 54 isconnected directly to the hub 50. Damper housing 56 is connecteddirectly to crankshaft drive plate 58. The hub of drive plate 58 issecured directly, preferably by bolts 60, to the end of crankshaft 14.

Damper springs 62 are situated between the damper drive plate 54 and thedamper housing 56. They are arranged in a circular fashion about theaxis of the hub 50, thereby providing a resilient connection between thecrankshaft and the clutch element 48 in known fashion.

The torque input shaft for the transmission is shown at 64. Although thetransmission illustrated in FIG. 1 can be used in the hybrid powertrainof the invention, other gearing arrangements also can be used to provideforward driving torque flow paths and a reverse ratio torque flow path.The gearing arrangement for the embodiment of FIG. 1 is similar to thegearing arrangement illustrated in U.S. Pat. No. 4,938,097 (S. L.Pierce), which is assigned to the assignee of the present invention.

Torque input shaft 64 is connected through a drive spline 66 to clutchcylinder 68 for forward drive friction clutch 70. When clutch 70 isengaged, a driving connection is established between shaft 64 and sungear 72 of a first planetary gear unit 74. A ring gear 76 is connecteddriveably to carrier 78 of gear unit 80. Carrier 78 is adapted to bebraked by selectively engageable low-and-reverse brake 82.

A sun gear 84 is connected driveably to the transmission torque inputshaft 64 through a reverse clutch 86. The ring gear 88 of the gear unit80 is driveably connected to the carrier 79 for the gear unit 74.

A direct-drive clutch assembly 90 connects the torque input shaft 64with the clutch cylinder 68. Clutch cylinder 68 also is connected to thesun gear 72 through the clutch 70, as mentioned earlier. Clutch 90 alsoconnects the shaft 64 to the ring gear 76.

The ring gear 88 of gear unit 80 defines a sprocket for a chain drive,indicated generally at 92. The driven sprocket of the chain drive, shownat 94, is rotatably mounted in the transmission housing on the axis oftorque output shaft 96. It is connected driveably to a sun gear 98 offinal drive gearset 100. The ring gear 102 of the final drive gearset100 is connected directly to the transmission housing.

The carrier of the gearset 100 is connected to differential carrier 104of a differential gear unit 106. Differential pinions are carried by thecarrier 104. They mesh driveably with side gears carried by torqueoutput shaft 96 and a companion output shaft 110. Each output shaft isconnected to vehicle traction wheels through a universal joint assemblyand axle half-shafts, not shown.

A friction brake band 108 surrounds a brake drum that forms a part ofthe reverse clutch 86. The brake drum is connected driveably to sun gear84 of gear unit 80.

FIG. 1 b shows a clutch and brake engagement and release sequence toestablish four forward-driving ratios and a single reverse ratio. Theclutches and brakes are identified in FIG. 1 b by the symbols RC, FC,DC, L/R and 2/4, which indicate, respectively, the reverse clutch 86,the forward clutch 70, the direct clutch 90, the low-and-reverse brake82 and the brake band 108. The symbols R, S and P (with appropriatesubscripts) in FIG. 1 a identify the ring gears, the sun gears and theplanetary pinion carriers, respectively.

To establish the first gear ratio in the forward-drive range, theforward clutch FC and the low-and-reverse brake L/R are engaged. Theforward clutch remains applied for operation in each of the first threeforward-driving ratios.

A ratio change to the second forward-driving ratio is obtained byapplying brake band 2/4 and releasing brake L/R. An upshift to the thirdratio is achieved by releasing brake band 2/4 and applying clutch DC.Fourth ratio, which is the top forward-drive ratio, is achieved byreleasing the forward clutch FC and applying reaction brake 2/4.

Reverse drive is obtained by simultaneously applying reverse clutch RCand low-and-reverse brake L/R.

An alternate embodiment of the transmission is illustrated in partialcross-sectional form in FIG. 2. A principal difference between thedesign of FIG. 2 with respect to the design of FIG. 1 is the lack of awet clutch in the design in FIG. 2 comparable to the wet clutch 46 ofthe design of FIG. 1. Because of the similarities in the design, thegearing system of FIG. 2 has been omitted in view of the completedescription herein of FIG. 1. Common elements of the design of FIG. 2with respect to the design of FIG. 1 have been indicated by the samereference numerals, although prime notations are used with the numeralsused in FIG. 2.

In both of the designs of FIGS. 1 and 2, the torque input shaft of thetransmission is connected mechanically to the engine crankshaft througha damper assembly. This differs from a conventional transmission whereinthe connection between the gearing torque input shaft and the enginecrankshaft is established by a hydrokinetic torque converter.

FIG. 3 is a schematic diagram of the overall transmission andstarter/alternator control system. This control system is dedicated tothe control of the transmission and starter/alternator, but it does notillustrate the vehicle system controller previously mentioned. Beforethe start of the shift, which is controlled by the control systemschematic illustrated in FIG. 3, the vehicle system controller commandsthe torque level for the motor based on the desired regenerative brakingtorque during a coasting downshift. Having determined the motor torquelevel, the control system of FIG. 3 proceeds to control power-onupshifts and coasting downshifts throughout the 5 modes previouslydescribed. At the end of a coasting downshift, the control of the motoris returned to the vehicle system controller.

FIG. 3 schematically illustrates the internal combustion engine at 10. Ahigh voltage battery 112 acts as a storage medium for storing electricalenergy generated by the starter/alternator 114. Torque developed by thestarter/alternator is distributed to a summing point 116, where it isalgebraically combined with engine torque to develop a transmissioninput torque at the mechanical torque flow path 118.

The multiple-ratio transmission for the hybrid electric vehiclepowertrain is shown at 12. It includes servo-operated friction clutchesand brakes, as previously described with reference to FIGS. 1 and 2. Itincludes also a variable force solenoid pressure control 120, which issupplied by a pressure controller associated with a central processingunit (CPU) 122.

Speed sensors 124 in the transmission 12 measure transmission inputspeed and transmission output speed. Those speed values are transferredthrough signal flow paths 126 and 128, respectively, to the CPU 122.

The CPU 122 is part of a microprocessor 130, which includes a read-onlymemory (ROM)132 containing a calibration table with which an eventtrigger can be obtained as a function of input torque. A servo-positionmeasurement mechanism can form a part of the transmission 12, as shownat 134, to effect a servo-position voltage signal at 136. The voltagesignal can be an indicator of 2/4brake torque capacity and may be usedto develop an event trigger voltage at 138 to establish the beginningand end of shift modes. An example of a device for sensing frictionelement torque capacity, based upon servo-position measurement, can beseen by referring to U.S. Pat. No. 6,110,068, which is owned by theassignee of the present invention. Other mechanisms, however, could alsobe used for establishing a trigger voltage at 138.

The microprocessor 130 includes a memory portion that contains astarter/alternator torque multiplier as a function of time. This isindicated at 140. The CPU uses the information at 140 to establish astarter/alternator torque multiplier signal at 142.

An accelerator pedal position sensor illustrated at 144 develops anaccelerator pedal position signal in signal flow path 146, which isreceived by ROM calibration table 148 in microprocessor 130. Acceleratorpedal position may use the calibration table at 148 to develop astarter/alternator base torque signal at 150. In the alternative, astarter/alternator base torque signal can be determined usingtransmission input speed as a variable.

The starter/alternator base torque signal at 150 is multiplied by themultiplier at 142 to develop a starter/alternator torque command at 152,which is distributed to the starter/alternator 114. Starter/alternatortorque, as previously mentioned, is used to develop transmission inputtorque. Transmission output torque for driving the vehicle tractionwheels is shown at 154.

Although FIG. 3 illustrates a transmission controller that is separatefrom the vehicle system controller, an integrated controller for theengine, the transmission and other vehicle systems could be used if thatis desired.

Various computer readable media, including random access memory,read-only memory and functional software instructions, as well as themanner in which information is stored in the media may be implemented,are well known in the art. The various functions are implemented by theCPU using the stored instructions or algorithms, which are executedrepetitively by the microprocessor in known fashion.

Strategy for Implementing Torque Modulation During a Power-On Upshift

The strategy for achieving a smooth power-on upshift is an open loopstrategy that uses event-based triggers to change, start, or stop thevarious events during execution of a power-on upshift. FIG. 4 a shows avariation of speed ratio during a power-on upshift event. FIG. 4 b showsthe starter/alternator torque multiplier, the information for which isretrievable from memory 140 of the microprocessor 130.

The output torque is shown in FIG. 4 c plotted against time. Thestarter/alternator provides a significant torque with a relatively quickresponse time.

The torque capability for the motor is plotted in FIG. 5 for a typicalembodiment of the invention. In FIG. 5, three different power levels areplotted for illustration purposes. A motor torque plot for a 20 kw motoris shown at 154. The corresponding plot for a 25 kw motor is shown at156. A 35 kw motor plot is shown at 158.

There are three distinct portions of a power-on upshift strategy inwhich the starter/alternator can be used to improve shift quality. Thetorque multiplier has a high value during the torque phase of the shift,as shown at 160 in FIG. 4 b. During the torque phase, the torque ratioacross the transmission changes from a low gear to a higher gear. Duringthis process, the output torque reduces to the next torque ratio timesthe input torque. This defines the start of a torque hole. This isindicated in FIG. 4 c at 162.

During the torque phase at 160, the strategy of the invention willreduce the severity of the torque hole 160, as shown at 164 in FIG. 4 c.This is done by increasing the input torque commanded using the torquemultiplier value at 160. The plot at 162 shows the variation of outputtorque when the strategy of the invention is not used. The improvementis apparent when the decrease in the torque level at 164 is compared tothe decrease in the torque level at 162. The torque disturbance(fluctuations) or shuffle at the output shaft thus is significantly lessnoticeable to the driver.

The starter/alternator command is set to zero at the termination of thetorque phase at point 166.

In implementing this control strategy, the start of the torque phasemust be detected. This can be done by using a position sensor 134described with reference to FIG. 3. The torque phase start can bedetected also using a so-called Kalman filter-based accelerationtrigger.

The output torque is reduced to the second gear level. The engine speedthen is pulled down toward the second gear speed. This results ininertia torque that defines the end of the torque hole shown at 162. Anacceleration trigger can be used to reduce the effects of the torquehole by releasing the off-going element at the appropriate time so thatthe amplitude of the torque disturbance will be minimized.

During the inertia phase, it is usual practice to reduce the inputtorque of a powertrain by retarding the spark of the engine, aspreviously explained. The starter/alternator makes it unnecessary toresort to a spark reduction method to achieve input torque reduction.During the inertia phase, the engine torque is not altered.

The total transmission input torque is the sum of the engine torque andthe starter/alternator torque. The starter/alternator torque iscommanded to absorb torque to reduce the total transmission input torquelevel during the inertia phase. When the inertia phase has begun, thecommanded starter/alternator torque is set to a calibrated value, which,as previously explained, can be a function of accelerator pedal positionor input speed. This value is multiplied by the torque multiplier ofFIG. 4 b, which provides shaping of the starter/alternator torquecommand to ease back the transition of the starter/alternator torque,before the end of the shift, to where it was at the start of the shift.Using starter/alternator torque modulation in this fashion, rather thana spark retard on the engine, improves the efficiency of the powertrainby storing energy in the vehicle battery 112 instead of wasting it asheat energy.

At the end of the shift, the oncoming clutch stops slipping, whichcauses an abrupt change in the kinematic state of the transmission. Thetorque carried by the clutch changes instantly when the clutch stopsslipping so that the value of the torque is significantly greater thanthe torque carried by a slipping clutch, which is a function ofpressure. A quick change in clutch torque can cause a torque shuffle inthe output shaft, as demonstrated in FIG. 4 c. This is due to the lowdamping rate of the driveline. To avoid this shuffle, the input torqueat the end of the shift is increased as the value of the torquemultiplier is increased. This is shown at 168 in FIG. 4 b. Therefore,the clutch torque at the end of the shift is more closely matched withthe slipping clutch torque capacity, thus minimizing the torque impulseinto the driveline.

The starter/alternator torque, as seen in FIG. 4 b, then is filteredback to the desired level as requested by the vehicle system controller.

FIG. 6 shows a time plot of normalized output torque and speed ratiovalues. The total transmission input torque is the combination of thestarter/alternator torque plus the engine torque. Although there may besome delay in effecting torque modulation using starter/alternatortorque and combining that torque through a control network, as indicatedin FIG. 3, the response time for motor torque control is well within therequirements needed to obtain shift quality targets. Further,repeatability of the torque modulation strategy is better than that foran engine fuel control strategy and is generally comparable to astrategy that uses a controlled engine spark retard.

In FIG. 6, torque fluctuations in the torque output of the transmission,without torque modulation, are plotted at 172. The torque outputfluctuations in the output shaft, when torque modulation is used, areplotted at 174. The speed ratio at 176 represents the levelcorresponding to a low gear ratio and the speed ratio at 178 correspondsto the upshifted value.

FIG. 6 demonstrates further that the inertia phase torque at 180 islower than the corresponding torque for the plot for a system that doesnot include torque modulation.

It is the combination of the hydraulic pressure control and the motortorque control that effects a smoothing and a shaping of the outputtorque during power-on upshifts. As mentioned previously, the shapedoutput torque reduces torque shuffle at the output shaft. To schedulethe computations and the output commands, the power-on upshift isseparated into 5 modes. These are illustrated in FIG. 7 a. Theinitialization mode (Mode 0), shown at 182 in FIG. 7 a, is triggeredwhen the desired gear is not the same as the value of the current gear.Variables are initialized in Mode 0 for the start of the shift.

Mode 1, shown at 184 in FIG. 7 a, is entered during the next mid-groundloop execution of the transmission controller 130, shown in FIG. 3. Mode1 lasts for a predetermined fixed time and is used to prepare theelements for pressure control. The friction element actuators are filledduring Mode 1 by boosting actuator pressure to a high initial value, asshown at 186 in FIG. 7 c. The pressure then is reduced sharply, as shownin FIG. 7 c, to a value 188, which is a function of input speed andtorque, from a table in memory.

Following the filling of the actuator for the oncoming friction element,the torque phase is entered at Mode 2, seen at 190 in FIG. 7 a. Thisoccurs after a fixed time for Mode 1 execution has expired. In Mode 2,the oncoming control pressure is commanded to ramp downward, as seen at192 in FIG. 7 c. The ramping occurs until the initial shift pressure isreached. In this way, the rate of pressure increase for the oncomingfriction element is controlled. The ramping ends as the torque phase isterminated, as shown at 194 in FIG. 7 c. The triggers, one of which isshown at 138 in FIG. 3, are monitored to indicate when the oncomingelement has sufficient capacity to allow the release of the off-goingelement. At this point, the off-going pressure is commanded to apredetermined clutch stroke value.

Mode 2, at 190 in FIG. 7 a, is complete after the percent ratio changehas exceeded a predetermined value or a set time has expired. A percentshift complete is computed by the CPU 122, as indicated at 196 in FIG.3.

The inertia phase, which is Mode 3 seen at 198 in FIG. 7 a, contains themajor portion of the speed ratio change time interval. In Mode 3, a PIDcontroller, based on comparison of the desired oncoming slip and theactual oncoming slip, generates the oncoming pressure command. This PIDcontrol of pressure is indicated in FIG. 3 as a function of the CPU 122.This closed loop pressure control is calibrated so that the shift rateinitially is high. It is decreased as the shift pressure is reducedduring the inertia phase, shown at 198 of FIG. 7 c. The reduced torquecapacity and the resulting reduced deceleration is used to reduce theseverity of shift shock that can cause shuffle after the shift iscompleted.

Mode 4, shown at 200 in FIG. 7 a, is entered when the percentage shiftcomplete exceeds a predetermined value. In Mode 4, the pressure isincreased to the value required for non-shifting operation. This is seenin FIG. 7 c at 202. This increase is achieved by an open loop control ofthe pressure command, as seen at 204 in FIG. 7 c.

Another variable that is controlled during a power-on upshift, asmentioned previously, is the transmission input torque. The motor basedtorque modulation strategy of the invention uses the event-basedtriggers to change a start or a stop of motor torque. The response ofthe motor when a command for torque is made, is a relatively quickresponse.

In Mode 1, the maximum value of torque the motor will be commanded toabsorb is determined as a function of pedal command. In Mode 2, themotor is commanded to start battery charging for absorbing torque afterthe start of the speed change is sensed.

In FIG. 7 a, output torque (axle half-shaft torque) is plotted at 206throughout the shift event. The engine speed is plotted at 208.

In FIG. 7 c, the commanded pressure at 204 results in a measuredfriction element pressure at 202.

When a servo-position signal is used, as shown at 136 in FIG. 3, a plotshown at 210 in FIG. 7 c is created. This plot will be an indicator ofwhen to start using starter/alternator torque to modulate the outputtorque. The beginning of the inertia phase can be anticipated as theslope of plot 210 is rising.

FIG. 7 b shows a shaping function, which is used during the shift event.This function is plotted in FIG. 7 b at 212. As indicated earlier, thisshaping function is used to determine the actual value of torque thatthe motor is commanded to absorb. Thus, the motor torque used to chargethe battery, which is plotted at 214 in FIG. 7 b, is changed, as theshift event proceeds, by the magnitude of the shaping function. Theshaping function information, as previously mentioned, is stored inmemory and called out by the program counter of the CPU and used as atorque multiplier in calculating motor torque as engine torque is addedto motor torque to obtain instantaneous, repetitive, input torquevalues.

In Mode 3, as the shift progresses, the starter/alternator is commandedto absorb less torque as shown at 216 in FIG. 7 b. This restores thetransmission input torque plotted at 218 in FIG. 7 b to its originalvalue. When the shift is completed, the increasing input torque reducesthe shift rate at the end of the shift further mitigating the shuffleproblem, previously described.

Strategy for Implementing Torque Modulation During Coasting Downshifts

The strategy for a coasting downshift is illustrated in FIGS. 8 a, 8 band 8 c, which are plots, respectively, of motor torque, pressurecommand, and input speed during a coasting downshift event. Before thestart of the coasting downshift, the vehicle system controller commandsthe torque level of the motor based on the desired regenerative brakingtorque. The beginning motor torque level is indicated at 220 in FIG. 8 aat the beginning of the shift. Torque control of the motor then istransferred to the transmission controller, indicated in FIG. 3, so thatthe shift event is coordinated with friction element control in thetransmission 12.

The oncoming clutch pressure is commanded by the transmission controllerto a maximum level to fill the oncoming clutch as quickly as possible,as shown at 222 in FIG. 8 b. The off-going pressure is commanded to alevel that is a function of input torque, as indicated in 224 in FIG. 8b. This pressure is high enough to hold the clutch just above itscapacity level. Then the off-going clutch pressure is ramped down, asshown at 226 in FIG. 8 b.

Motor torque is ramped from the current negative level to a lowernegative level, as shown at 228 in FIG. 8 a. That level is calibratedfor each shift. Reducing the amount of negative regenerative brakingtorque in this way reduces the effort needed to bring the transmissionto the downshifted gear, thus making the transition from one gear to theother more smooth. Motor torque is ramped to the new level at a ratethat depends upon whether the engine is connected.

The boost mode at 222 in FIG. 8 b is time-based. During the start mode230, which is Mode 2 seen in FIG. 8 a, the oncoming pressure iscommanded to an open loop pressure 232, seen in FIG. 8 b. That pressurecommand is increased if the coasting downshift is made by manuallydownshifting the transmission, or if the vehicle wheel brakes areapplied.

During a downshift, it is expected that the transmission input speedwill increase. If the off-going element losses capacity before theoncoming element is capable of holding input torque, the input speedwill decrease due to the negative torque load of the motor. A reductionin input speed is an indicator of off-going friction element slip. Theoff-going friction element pressure ramp continues until off-goingfriction element slip is detected or oncoming capacity is detected.

The motor torque ramp, shown at 234, continues through the boost mode,the start mode and the inertia mode, as shown at 236 in FIG. a, untilthe desired torque is reached, as shown at 228 in FIG. 8 a. It is heldat that level until it is requested to return beginning at point 238 inFIG. 8 a. The level of motor torque is a calibratable function for eachshift.

The end of the start mode is detected through detection of oncomingclutch capacity. When the oncoming clutch has capacity, the transmissioninput speed direction will begin to move toward the speed at thedownshifted speed ratio. This is seen in FIG. 8 c at 239. Once acalibratable percentage of shift is completed or a time-out occurs, thestart mode is completed.

In the inertia mode, oncoming pressure is commanded through a closedloop PID controller to follow a speed profile to the downshifted gear.This is indicated at 240 in FIG. 8 b. There is a minimum clip on thepressure command, as shown at 242 in FIG. 8 b, which would be raised tomake the coast shift more aggressive under two conditions: first, duringa manual downshift, an aggressive feel is desired so that the vehiclenoticeably slows in response to the driver's desire for hill braking;second, under fast vehicle deceleration, the shift needs to be completedfaster and a more aggressive shift rate is acceptable.

At the beginning of the start mode, the off-going clutch is commandedoff. The motor torque is expected to be at a level needed to completethe ratio change. As the shift advances toward completion, the motortorque is decreased, as shown at 244 in FIG. 8 a, toward the originaltorque level that was commanded at the start of the shift. The point 238at which the down slope is begun is a function of percentage shiftcomplete. To make the down slope responsive to vehicle decelerationrate, the down slope rate is a function of percentage shift complete.

The inertia mode is exited when the percentage shift complete is near100%. In the end mode 246 in FIG. 8 a, the oncoming pressure command isincreased through a parabolic filter to the maximum command. This isseen at 248 in FIG. 8 b. At the completion of the shift, motor controlis returned to the vehicle system controller.

FIGS. 9, 10 and 11 illustrate data for an actual 4-3 coasting downshift.In order to correlate the events that are illustrated in FIGS. 9, 10 and11 with the shift modes illustrated in FIGS. 8 a, 8 b and 8 c, the plotsof FIGS. 9, 10 and 11 include mode plots 250, 250′ and 250″,respectively. FIG. 9 shows a motor torque profile, which linearlyincreases, as shown at 252. Torque is requested to increase throughoutthe first two modes and does not reach its steady-state value until theshift is near completion at 254. Mid-way through the inertia phase, themotor torque is brought back down near the value needed for thedownshifted ratio as seen at 256. The output torque, before and afterthe coasting downshift, is at approximately the same level with minimumdisturbance during the inertia phase. This is shown at 258.

FIG. 10 shows the profile of the oncoming commanded pressure at 260 andthe off-going commanded pressure at 262, as well as the actual pressures264 and 266 corresponding, respectively, to the commanded pressures 260and 262. The off-going friction element in the case of FIGS. 9, 10 and11 is the 2-4 shift command. The off-going friction element pressurecommand is set to zero during the boost phase, as seen at 268 in FIG.10, in order to speed its response. The forward clutch is the oncomingelement in the case of a 4-3 coasting downshift. Its response is fast,so less boost is needed.

FIG. 11 shows the change in input speed for this 4-3 coasting downshift.Input speed varies smoothly from the fourth ratio level to the thirdratio level, as seen at 270.

Although an embodiment of the invention has been described, it will beapparent to persons skilled in the art that modifications may be madewithout departing from the scope of the invention. All suchmodifications and equivalents thereof are intended to be defined by thefollowing claims.

1. A method for controlling engagement and release of pressure-actuatedtorque establishing friction elements during a power-on upshift of amultiple-ratio transmission in a hybrid electric vehicle powertrainhaving an engine, an electric motor and a battery, the motor beingdisposed in a power flow path between the engine and a power inputelement of the transmission as driving power is delivered to atransmission power output member, and an electronic controllerresponsive to powertrain operating variables for controlling actuatingpressure for the friction elements by releasing actuating pressure foran off-going friction element in synchronism with increasing actuatingpressure for an oncoming friction element during a power-on upshiftevent, the method comprising the steps of: increasing oncoming frictionelement actuating pressure at the beginning of the power-on upshiftevent; reducing actuating pressure for the off-going friction elementduring a torque phase of the power-on upshift; increasing electric motortorque during the torque phase; and reducing electric motor torqueduring an inertia phase of a power-on upshift whereby torque variationsin the transmission power output member are modulated.
 2. The method setforth in claim 1 wherein the step of increasing electric motor torquecomprises obtaining a torque multiplier stored in controller memoryduring each control loop of the electronic controller; and multiplyingmotor torque by the multiplier during the torque phase to increaseeffective motor torque during the torque phase.
 3. The method set forthin claim 1 wherein the step of reducing electric motor torque during theinertia phase comprises obtaining a torque multiplier stored incontroller memory during each control loop of the electronic controller;and multiplying motor torque during the inertia phase to decreaseeffective motor torque during the inertia phase.
 4. The method set forthin claim 2 wherein an effective torque in the transmission power outputmember at the end of the power-on upshift is the algebraic sum of motortorque and engine torque multiplied by instantaneous transmission speedratio.
 5. The method set forth in claim 3 wherein an effective torque inthe transmission power output member at the end of the power-on upshiftis the algebraic sum of motor torque and engine torque multiplied byinstantaneous transmission speed ratio.
 6. The method set forth in claim1 including the steps of determining a time during a power-on upshiftevent when the oncoming friction element achieves torque establishingcapability; and controlling oncoming friction element pressure duringthe inertia phase using a commanded closed loop pressure.
 7. The methodset forth in claim 6 including the step of detecting the end of theinertia phase; and increasing the commanded pressure on the oncomingfriction element to effect full oncoming clutch capacity at the end of apower-on upshift event.
 8. The method set forth in claim 7 wherein thestep of increasing the commanded pressure on the oncoming frictionelement occurs as the oncoming clutch becomes fully engaged followingslipping of the oncoming clutch whereby an effect of a change inoncoming friction element dynamics is modified thus reducing torquefluctuation in the torque output member.
 9. The method set forth inclaim 8 including a step of increasing the commanded pressure on theoncoming friction element prior to the torque phase followed by a returnof the commanded pressure on the oncoming friction element to an initialpressure value at the start of a power-on upshift event.
 10. A controlsystem for controlling ratio changes in a multiple-ratio transmission ina hybrid electric vehicle powertrain, the transmission havingselectively-engageable, pressure-actuated torque establishing frictionelements for effecting powertrain upshifts; the powertrain having anengine, an electric motor and a battery, the motor being disposed in apower flow path between the engine and a power input member of thetransmission as driving power is delivered to a power output member; thecontrol system comprising an electronic controller responsive topowertrain operating variables for controlling actuating pressure forthe friction elements; the control system controller being configured tocontrol actuating pressure for the friction elements by releasingactuating pressure for an off-going friction element in synchronism withincreasing actuating pressure for an oncoming friction element during apower-on upshift; the motor being disposed in series relationship withrespect to the engine and the transmission; and the control systemincluding means for increasing motor torque during a torque phase andreducing motor torque during an inertia phase of a power-on upshift ofthe transmission.